Building Performance With District Cooling

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1 Reprinting this proof for distribution or posting on web sites is not permitted. Authors may request permission to reprint or post on their web site once the fi nal version has been published. A reprint permission form may be found at Building Performance With District Cooling By Eric M. Moe L ow chilled water temperature differential ( T) T continues to be a major district cooling weakness. Poor T T performance at cooling coils leads to lost cooling capacity, wasted energy, extra cost and added complexity for a thermal utility, its chilled water customers, or both. some efforts to reduce the flow simply shift issues from the central plant to the building. If the approach to raising T reduces cooling coil capacity, increases building pump and fan energy consumption, or adds complexity, the benefit of lower chilled water rates for chilled water customers is marginal at best. Rapid district cooling industry growth depends on the ability to satisfy existing customers and effectively separate the advantages of district energy from selfgenerated heating and cooling. Choosing an approach that optimizes performance at cooling coils can simplify building operation, reduce capital expenses, increase system reliability, improve diversity, and consume less (total) energy. In a large chilled water system, the root cause of low T T is typically within your customer s building at the load. Cooling coil and control valve performance must be considered to fully address systemwide performance issues. To encourage customers to invest in technology to improve T T performance in their buildings, an increasing number of utilities have established chilled water rates that vary inversely with T. T Figure 1 is an example of rates charged to customers from one prominent university in the United States. As you can see, the lower the T, T the greater the rate. Conversely, customers that minimize their flow rate per ton cooling are rewarded. Customers want to keep their occupants comfortable, minimize total energy cost, and reduce the equipment required for district cooling service. Unfortunately, Chilled Water Customer Interconnections Figure 2 represents a variety of customer building connections to the chilled water distribution. Decoupled buildings are designed to prevent performance in one building from adversely affecting the system. Direct connect strategies prevent return water from blending with supply at the building. However, without high T T performance at the coils, these About the Author Eric M. Moe is director of business development at Flow Control Industries in Woodinville, Wash. 4 6 A S H R A E J o u r n a l a s h r a e. o r g J u l y

2 strategies may return water to the plant at less than design temperature. To achieve excellent performance in the building, distribution, and central plant, it is critical for cooling coils to generate high T T through the range of load conditions. Furthermore, chilled water supply temperature (CHWST) shouldn t rise to a level that compromises coil capacity, causes humidity problems, or creates a net rise in the total cooling energy consumed. Direct connected building with booster pump: adds head to the supply to ensure flow through coils; it may take flow from or create reverse flow at other buildings. Direct connected building with pressure knockdown valve: reduces the inlet pressure to coils and control valves. Decoupled building with both supply temperature and differential pressure control: uses the valve in the chilled water supply line to regulate the inlet pressure and the valve in the crossover bridge to regulate an increase in the CHWST. Decoupled building with return water temperature control: circulates chilled water in the building until the return water temperature rises high enough to return it to the plant, which is intended to compensate for poor T T performance across coils. Decoupled building with heat exchanger: completely separates the building from the distribution flow; it creates additional heat transfer loss. Direct connected building: simplest approach requiring the least amount of equipment and physical space in the build- ing; it is susceptible to changes in pressure with conventional control valves. Direct connected building with booster pump and bypass: simple approach boosts the supply pressure but only if required at high load; it is susceptible to changes in differential pressure with conventional control valves. Decoupled Building with Return Water Temperature Control Return water temperature control is a common strategy used in buildings served by large district cooling systems. By decoupling the building flow from the primary or secondary distribution flow, chilled water is recirculated in buildings until enough heat has been gained to produce adequate T T (Figure 2 [Scheme D] and Figure 3). While this technique has been used as an intermediate step to generate the desired temperature rise at the plant or through the main distribution pipes, it is fraught with potential problems for customers. When return water is permitted to blend with supply, CHWST increases. This commonly leads to building performance issues including wasted pump and fan energy, lost coil capacity, and poor humidity control. On a system level, pumps used in buildings close to the plant may starve buildings downstream of the pressure and flow required to serve the load. Hydraulic gradient analysis is one technique that may be used to evaluate system design and operation. 1 System pressure typically is plotted along the vertical axis and distance from J u l y A S H R A E J o u r n a l 4 7

3 the central plant plotted along the horizontal axis. Pumps develop head. Chilled water flows from high to low pressure. In the hydraulic gradient shown in Figure 3, plenty of distribution head exists to serve the load without running the building pumps close to the plant. 2 The following describes how each building in the system is operating. Building A: a portion of the chilled water return blends with supply in an effort to reach a 54 F (12.2 C) return water temperature setpoint. CHWST to the cooling coils increases. Building B: a portion of the supply water bypasses coils in the building when the return water temperature leaving the coils exceeds 54 F (12.2 C). Building C: no chilled water flows through the crossover bridge. It is easy to see from Figure 3 that the return water temperature control valves for the buildings consume all the remaining head developed by the building pumps, as well as any additional pressure between the supply and return header. For the load condition shown, none of the pumps in Buildings A, B and C must be run. When all building pumps are run, as shown in Figure 3, the suction head pressure differential in the distribution steps down as it passes each additional building on the loop. This process can tend to starve buildings downstream of the pressure differential and flow that they need. Larger building pumps that generate more suction head are not the solution to generate chilled water flow in starved buildings. Factors Influencing Coil and Control Valve Performance Through the cooling load range, the T T performance of a properly selected coil can decline for a number of reasons. 90 psi Distribution Head 30 psi 40 F Pump Suction Primary or Secondary Pumps 54 F 105 psi Bldg. A 45 F Return Water Blends With Supply 54 F 40 F Supply Water Blends With Return 54 F Bldg. B No Flow Through Bridge Rate ($ per ton/h) Chilled Water T T ( F) Figure 1: Incentive chilled water rates. gpm 24 = ton T A B C D E F G Figure 2: Common chilled water interconnections to buildings. Bldg. C Return Water Temp. Control Valves Distance from Central Plant 20 psid 15 psid 50 psi Dirty fins and coils, high chilled water supply temperature, low supply air temperature, improper piping, over-pumping, poorly modulating control valves, improper valve sizing, and control valve hunting are a few significant factors that will adversely affect chilled water T T performance. A significant body of published literature discusses low T problems as well as design, operation, maintenance and control strategies typically used to attempt to mitigate the issues. Despite the industry s best efforts, low T T continues to be an un- resolved problem in the majority of large facilities served with central cooling equipment. Although control valve selection and performance is a key contributor to total system performance in operation, it seldom receives the attention it deserves. Chilled Water Supply Temperature Figure 4 illustrates how T, T flow, and capacity of a 45/55 split ARI-certified cooling coil varies with a change in the chilled water supply temperature. At 100% design flow for the coil, observe the dramatic reduction in capacity when % Total Capacity 140% 120% 100% 80% 60% 40% CHWST 39 F 45 F 42 F 48 F 20% 20 F, 15 F, 10 F T 0% 0% 20% 40% 60% 80% 100% 120% % Design Flow Figure 3 (left): Return water temperature control. Figure 4 (right): Effect of rising CHWST on coil capacity, flow rate and T. 4 8 A S H R A E J o u r n a l a s h r a e. o r g J u l y

4 Pump Discharge Loss Chiller Loss Pump Head Building 1 System Pressure (ft) Closest Circuits Furthest Circuit Building Differential Pressure (psid) Building 2 Pump Suction Loss 0 50 One Week (Summer 2003) Distance Figure 5 (left): Variable primary flow hydraulic gradient. Figure 6 (right): Typical differential pressure at two buildings on university campus. the CHWST rises to 48 F (8.9 C). Also, observe the high T T that the coil is capable of producing when low CHWST is maintained. Proper Control Valve Sizing/Rules of Thumb The hydraulic gradient in Figure 5 is used to illustrate a simple variable primary flow system with direct return piping. As shown, the closer a circuit is to the plant, the greater the differential pressure that is consumed by the coil and control valve between the supply and return header. ASHRAE is reevaluating recommendations regarding control valve selection. 3 None of the current recommendations take into account dynamic system performance Circuit Near Plant or consider future variations in the hydraulic profile as systems expand. Control valves are commonly sized for peak load conditions using rules of thumb without taking into consideration the differential pressure at each circuit. Some of the most common methods for control valve selection this author has seen used are: Size all control valves for 5 psi (~12 ft [~36 kpa]) pressure drop; Select control valves for the same pressure drop as the coil it serves; Most Remote Circuit Total Pressure Drop 96 ft 20 ft Coil Pressure Drop 8 ft 8 ft Valve Pressure Drop 88 ft 12 ft Coil Design Flow 225 gpm 225 gpm Wide Open Valve gpm gpm Percent Flow Through Valve 36.9% 100% At Coil Design Flow Table 1: Impact of sizing all control valves for 12 ft pressure drop. Select control valves for line size or one smaller; Don t choose control valves smaller than half the line size; Choose control valves for 25% 50% of the system pressure; and Ensure the valve authority is relative to the pressure drop in the circuit served. Using Figure 5, consider the impact of using the same pressure drop to size each coil control valve when design flow for Total Load T T CHW Flow Dist. Drop Coil Drop Valve Drop Total PD Pump Power Motor Power 350 tons 18.0 F 467 gpm 1.22 ft 1.39 ft ft 14.6 ft 2.0 bhp 1.7 kw 350 tons 15.0 F 560 gpm 1.75 ft 2.00 ft ft 15.8 ft 2.6 bhp 2.2 kw 350 tons 12.0 F 700 gpm 2.73 ft 3.13 ft ft 17.9 ft 3.7 bhp 3.1 kw 350 tons 7.5 F 1120 gpm 7.00 ft 8.00 ft ft 27.0 ft 9.0 bhp 7.5 kw Table 2: Building pump motor power requirements with decreasing T. Sensible Load Exhaust Air Supply Air Airflow Dist. Drop SP Setpoint Total PD Fan Power Motor Power 280 tons 75 F 55.0 F 154,839 cfm 2.08 in. H 2 O 1.00 in. H 2 O 3.1 in. H 2 O bhp 83.5 kw 280 tons 75 F 60.0 F 206,452 cfm 3.70 in. H 2 O 1.00 in. H 2 O 4.7 in. H 2 O bhp kw 280 tons 75 F 62.0 F 238,213 cfm 4.93 in. H 2 O 1.00 in. H 2 O 5.9 in. H 2 O bhp kw 280 tons 75 F 62.1 F 240,000 cfm 5.00 in. H 2 O 1.00 in. H 2 O 6.0 in. H 2 O bhp kw Table 3: Fan cfm and motor power requirements with rising SAT. J u l y A S H R A E J o u r n a l 4 9

5 System Pressure (ft) Low 6 Low 8 Design ft drop across coils and control valves 60 Expansion Tank Pressure 40 Distance Figure 7: Hydraulic gradient at design through range of T performance. each coil is 225 gpm. It is easy to see how control valves close to the plant can become grossly oversized for the application. In Table 1, control valves were selected with a Cv of 100 to deliver at least 225 gpm (14.2 L/s) design flow assuming a 12 ft (36 kpa) pressure drop across each control valve. No consideration was given to the location of the valve in the system. At design flow through the coil, the control valve in the circuit near the plant is only flowing 36.9% of its capability. If this is a commercial quality control valve with typical construction and range, T T performance will suffer, especially at part load. Uncertainty in the Hydraulic Profile It often is not well understood what differential pressure to use to size conventional control valves. In many cases, the absence of this information leads to control valve sizing by rules of thumb. Figure 6 illustrates differential pressure data from two buildings on a university campus. Knowing the system has low T, T and the system is expanding, what is the best strategy to size control valves for these buildings? Even if conventional control valves are properly selected now, the hydraulic profile may change in the future. Control Valve Hunting If, at a specific cooling coil, the differential pressure changes across the control valve without a change in load, the flow changes immediately. In any large system, pressure fluctuations are unavoidable as pumps go on- and off-line and change speed, and as loads and flows vary at other locations. Figure 6 is a good example of this. It takes time for a conventional control valve actuator to return the valve to the right position for the load. Pressure dependent control valves hunt constantly as the hydraulic profile changes, even with sophisticated controls. Valve and Actuator Capability As Figure 7 illustrates, when T T is low, the differential pressure close to the plant is higher than anticipated. Many conventional control valves and their actuators lack the capability to handle Flow P 1 P 1 P 2 Control Shaft Rotates to Modulate Flow Flow Control Surfaces Springs Piston Flow Figure 8: Pressure independent modulating two-way control valve. P 3 Seal the extra differential pressure and may begin to operate like on-off valves. With the added pressure, balancing valves only function to clip maximum flow and do nothing to optimize coil performance at part load. Pressure Independent Chilled Water Control With pressure independent control, the flow rate at any given cooling coil changes only when the load changes. Typically, the air handler leaving air temperature relative to setpoint is used to control the actuator to position the valve. Figure 8 illustrates pressure independent modulating two-way control valves. Mechanically, the piston and spring in the valve function to maintain a constant differential pressure across the control surface within the valve (P1 P2). To change the flow rate, the stem must be rotated. A change in the pressure upstream or downstream of the valve will not change the flow. Applying pressure independent modulating two-way control valves eliminates the effect of pressure fluctuations on the flow rate through cooling coils. It also creates a system robust to uncertainty or changes in the hydraulic profile. High T T at all load conditions is the result. 4,5 An ideal building configuration ensures that high T is achieved at each coil in the building and eliminates the need for building pumps to be run when not required (Figure 2, Scheme G). If the building is not in a hydraulically remote location, the building pump can be removed so that only a supply and return pipe is necessary to provide chilled water (Figure 2, Scheme F). By using pressure independent modulating control valves and eliminating return water blending with supply, low CHWST is maintained, T T is maximized, and the coil s capacity is increased. It is easy to conceptualize how raising the T T and minimizing the use of building pumps, crossover bridges, and deny valves saves pump energy. Variable speed fan energy savings also may be realized by enabling the system to maintain low leaving air temperature off the coil. Energy Savings Estimate The following example is intended to illustrate the economic impact of the pump and fan energy savings possible in a single 5 0 A S H R A E J o u r n a l a s h r a e. o r g J u l y

6 building that converts from return water temperature control with conventional control valves to direct connect with high quality pressure independent modulating control valves. The example also factors in an incentive paid by the thermal utility for raising T T across the building. Figure 9 is an illustration of the conversion F 12 ft 12 ft Building Mechanical Configuration The building considered is a 100% outside air facility with three large variable speed air-handling units designed for 80,000 cfm ( L/s) and 6 in. H 2 O (1.5 kpa) pressure drop each. Cooling coils are designed for 42 F (5.5 C) CHWST and a 15 F (8.4 C) T. T The pump serving the building is capable of delivering 1,120 gpm (71 L/s) at 27 ft (81 kpa) of head. Design cooling load is 700 tons (2462 kw) (400 tons sensible [1407 kw]). The estimate considers the effect of rising supply air temperature and declining T T on a part load day. Building Pump Energy With a total design cooling load of 700 tons (2462 kw), design flow for the building at 15 F (8.4 C) T T is 1,120 gpm (71 L/s). A VFD controls pump speed to maintain 12 ft (36 kpa) across the hydraulically most remote coil control valve. T T = CHWRT CHWST gpm = tons 24 T = gpm PD bhp pump pump 3960 η kw pump_motor = η pump bhp pump η motor Actual Building Load 350 tons (1231 kw) total cooling load (280 tons sensible [985 kw]) 40 F (4 C): design CHWST 55 F (13 C): design CHWRT. Variable Speed Pump Details One 1,120 gpm (71 L/s) pump 7 ft (21 kpa) drop each through the supply and return (at design flow) 8 ft (24 kpa) drop across coil (at design flow) 12 ft (36 kpa) drop across control valve (always) 85% pump efficiency (assumed) 90% pump motor efficiency (assumed) Using the above equations, the pump design power is: bhp = 9 (6.7 kw). If the system pump is providing enough pressure, the best option is to serve the load without the building pump running. At locations closer to the central plant, the building pump is removed or a bypass is installed around the pump (Figure 2, 42 F 62 F LAT BEFORE 54 F T 55 F LAT AFTER 60 F Figure 9: Change from decoupled building with return water temperature control to directly connected building with pressure independent modulating two-way control valves. Schemes F and G). Alternatively, as shown in Table 2, if the building pump is run, the least energy consumption occurs at the highest T. T Note that with T T at 7.5 F (4.2 C) the pump motor is running at maximum power despite the installed VFD. Only by minimizing both the flow rate and pressure drop by raising T will the VFD perform as intended to save energy. If a conventional system is designed with circuit setters, these may limit total flow through the building to 1,120 gpm (71 L/s). However, if CHWST is rising due to return water blending with supply, coil capacity is reduced and it will need more flow to satisfy the same load. The resulting effect is a rise in supply air temperature and perhaps an increase in airside energy or reduction in occupant comfort. Locating the differential pressure sensor across the hydraulically most remote control valve has an impact on pump energy consumption. Since an installed cooling coil is merely a fixed pipe, the differential pressure can be placed across the hydraulically most removed valve (in lieu of the entire circuit) to save pump energy at part load. AHU Fan Energy For this portion of the analysis, we will look only at the air handler fan energy that is consumed at peak load in a typical 100% outside air building as the supply air temperature rises. cfm = 12,000 tons sensible (EAT LAT) = cfm actual bhp PD fan fan 6356 η kw fan_motor 42 F η fan = bhp fan η motor J u l y A S H R A E J o u r n a l 5 1

7 Temperature ( F) May 21, 2004 OAT SAT (Design 55 F) CHWST (Design 45 F) T (Design 12 F) Actual Building Load 350 tons (1231 kw) total cooling load (280 tons sensible [985 kw]) 75 F (23.9 C): space temperature setpoint 55 F (12.8 C): leaving air temperature setpoint. Variable Speed Fan Details Three 80,000 cfm ( L/s) air handlers 6 in. H 2 O (1.5 kpa) total pressure drop at 80,000 cfm ( L/s) 1 in. H 2 O (0.25 kpa) static pressure setpoint 75% fan efficiency (assumed) 90% fan motor efficiency (assumed) Using the above equations, the fan design power is: bhp = (75 kw) each. For the analysis below, the static pressure setpoint is 1 in. H 2 O with 6 in. H 2 O (1.5 kpa) total pressure drop at 80,000 cfm ( L/s) through each air handler. To simplify analysis, fan efficiency was assumed to be 75% and fan motor efficiency was assumed to be 90%. Table 3 illustrates airflow and motor power requirements in the building at part load with rising supply air temperature. In many chilled water system applications with return water temperature control, rising CHWST prevents the system from achieving low supply air temperature (SAT). The price paid is in additional fan horsepower and thermal comfort. In the case illustrated previously, if the SAT could be reduced from 62 F to 55 F (16.7 C to 12.8 C) without violating minimum airflow requirements, fan energy could be reduced from to 83.5 kw, a 65.8% reduction. Keys to Reduced Building Fan and Pump Energy It is the combination of low supply air temperature and high T T that will minimize pump and fan energy consumption in buildings. If return water is blended with supply in any application, coil capacity is reduced, T T declines, and gpm/ton rises. Variable speed fan energy also will rise if the intended SAT rises above setpoint. Figure 10 illustrates the adverse effect of return water temperature control in an actual system. In this laboratory, the CHWST to the building is 45 F (7.2 C), but it rises in the building to as high as 57 F (13.9 C). SAT rises to 62 F (16.7 C) Temperature ( F) June 30, 2002 Figure 10 (left): Cooling coil performance with return water temperature control. Figure 11 (right): Cooling coil performance with pressure independent modulating control valves. OAT SAT (Design 55 F) CHWST (Design 42 F) T (Design 15 F) resulting in excess airflow, humidity concerns and excess fan energy consumption. T T within the building is far below the 12 F (6.7 C) design T T for the coils. In contrast, Figure 11 illustrates performance in a system designed as shown in Figure 8. No return water blending with supply is permitted. Pressure independent modulating control valves on coils are used. CHWST remains at 40 F (4.4 C) permitting the SAT to be reduced to 55 F (12.8 C). Airflow in this case can be minimized to save fan energy. T T exceeds the 15 F (8.4 C) T T for the coil at all load conditions, permitting chillers in the plant to be fully loaded and pump energy reduced. Total cooling energy costs in a customer s facility are a combination of chilled water rates from the utility and the electricity consumed operating pumps and fans in the buildings. As an illustration, Figure 11 shows the performance improvement achieved when a decoupled building with return water temperature control is converted to a direct connect building with pressure independent modulating control valves at the cooling coils. Table 4 is a summary of expected customer energy cost savings for three 24/7 buildings with the design cooling characteristics described previously but located in three different cities. Electricity rates are assumed to be $0.07/kWh. Incentive chilled water rates, as described in Figure 1, are applied. Conclusions Return water temperature control is one example of a design strategy that increases T T across the building to reduce chilled water flow at the plant and in the main distribution. Unfortunately, the benefit to the central plant and distribution can compromise building performance. With the growing application of high quality pressure independent modulating control valves, no need exists to decouple buildings just to address low T T issues at coils. Despite unavoid- able system pressure fluctuations or the evolution in the hydraulic profile over time, these valves help generate the maximum possible T T to minimize flow requirements at any load. If the district cooling industry hopes to further improve performance over self-generated cooling, it is vital that we choose components that maximize chilled water T performance at cooling coils. Incentive chilled water rates help foster invest- 5 2 A S H R A E J o u r n a l a s h r a e. o r g J u l y

8 ments in equipment to optimize performance. High T in the distribution and at the central plant improves diversity and increases available system capacity permitting the system to cool more customers with less equipment and less energy. 6 This should translate into more cost-effective service. Typical benefits customers can expect with directly connected buildings and pressure independent modulating control valves include less fan energy consumption, less building equipment, simpler system operation, reduced capital and operating expenses, and better humidity control. References 1. Rishel, J System Optimization with Hydraulic Gradient Analysis and Pressure Independent Control Valves. University of Wisconsin Chilled Water Plant Seminar. 2. Moe, E.M Win-Win Control Strategies for District Cooling Customers. Presented at the International District Energy Association Campus Conference. 3. Control Valve Selection Sponsoring Technical Committee 1.4, Control Theory and Application. Project Work Statement Version 1, Feb Control valves exceed design engineers expectations The Chief Engineer (12). Seattle Boston Dallas 50% Load (tons) Sensible Load (tons) Equivalent Time at 50% Load (hours/year) 1,747 3,119 7,124 Pump Power Savings (kw) 7.5 F to 18 F T T Fan Power Savings (kw) 60 F to 55 F LAT Annual Electricity Savings (kwh) 158, , ,284 Electricity Consumption ($/kwh) $0.07 $0.07 $0.07 Electricity Savings (per year) $11,128 $19,868 $45,380 Total Chilled Water Cooling (ton-hours/year) 611,450 1,091,650 2,493,400 Incentive to Increase T At Building Interface from 10 F to >15 F $ $ $ Reduction in CHW Expense (per year) $28,188 $50,325 $114,946 Total Customer Energy Cost Savings (per year) $39,316 $70,193 $160,326 Table 4: Annual cooling energy cost savings for customers. 5. Borer, E. and Schwartz, J High Marks for Chilled-Water System: Princeton Upgrades and Expands. District Energy, First Quarter. 6. Skoglund, P.K Control Your Chilled Water Save Energy / Increase Capacity. International District Energy Association Conference Presentation. Advertisement formerly in this space. J u l y A S H R A E J o u r n a l 5 3